Elastic fluid compressor



Feb. 21, 1967 A GREGORY ELASTIC FLUID COMPRESSOR 3 Sheets-Sheet 1 Filed Dec. 20, 1963 INVENTOR ALFRED T. GREGORY BY ZWW 7 /740 ymwu f HIS ATTORNEYS 1967 A. T. GREGORY ELASTIC FLUID COMPRESSOR 5 Sheets-Sheet 2 Filed Dec. 20, 1963 l NVENTOR ALFRED T. GREGORY HIS ATTORNEYS Feb. 21, 1967 A. T. GREGORY 3,305,165

ELASTIC FLUID COMPRESSOR Filed Dec. 20, 1963 3 Sheets-Sheet 5 INVEN R g ALFRED T.GREG Y B 5M, 24. M Lu. 7

HIS ATTORNEYS United States-Patent O l 3,305,165 ELASTIC FLUID COMPRESSOR Alfred T. Gregory, West Hills, Huntington, N.Y. 11743 Filed Dec. 20, 1963, Ser. No. 332,158 7 Claims. or. 230-124 This invention relates to a novel and improved compressor for compressing elastic fluids, such as gases or mixtures of gases like air, which is particularly adapted for compressing fluids at small mass flow rates for which either axial flow or centrifugal types of compressors cannot be used efficiently. The invention, however, is not limited to compressors with low mass flow rates.

It is an object of this invention to provide a novel and improved compressor of relatively simple and compact structure and of small size for compressing air or other gases at low mass flow rates and relatively high efficiencies.

It is a further object of this invention to provide an improvement in the gas compression process, thus making it possible to construct relatively small compressors of simple structure, even for large mass flow rates, with high pressure ratios and efiiciencies.

Both the conventional axial fiow and centrifugal gas compression machines experience high blade-tip losses when blade heights are made very short, as is requisite for small mass fiow or low volume flow rates. The compressor blades, in a compressor in accordance with the invention, however, have no tip clearance so that tip losses are eliminated. Furthermore, the two-dimensional flow that exists in the radial in-fiow portion of the subject compressor avoids the losses due to turbulence which occur in presently known types of compressors with three-dimensional fiow characteristics. The combined effects of these and other factors provide greatly improved overall compressor performance, particularly at low mass or volume flow rates.

More particularly, in the compressor of the invention the gas flows in a radially inward direction from an annular plenum chamber into a rotating centripetal stage of thin airfoil section blades where it is compressed. It continues to flow radially inward through a row of stator blades and is then discharged in an axial direction into a centrifugal stage where it is further compressed and discharged outwardly.

The leading edges of the centripetal first stage rotor blades are at a constant radius from and are parallel to the axis of rotation of the rotor. The blades are straight in the axial direction and all blade elements in planes perpendicular to the axis are identical. The relative Mach number of the entering air is thus constant over the axial length of the blade, and the vector diagrams, blade solidity, and blade loading are the same for all blade elements. There is little or no change in the width of the air passage through the rotor blades, so that the flow remains substantially two-dimensional in character. Conditions are therefore ideal for the obtaining of optimum performance with transonic blades, which are accordingly preferred for the centripetal stage.

In contrast, the presently known transonic compressors have all been of the axial flow type. Vector diagrams vary greatly from root to tip of the blades and, while the entering Mach number at the tip is usually slightly supersonic, that at the root is subsonic, especially in the first blade rows Where hub-tip ratios are made small in order to obtain high mass flow rates for a given size machine. Because of the reduced blade velocity at the root, the blade turning angles at the root are much greater than those at the tip. Furthermore, since the blades have a large amount of twist, higher blade thickness ratios are required in the root sections, thereby further impairing blade performance in those areas, in order both to reduce 3,365,155 Patented Feb. 21, 1967 the amount of untwisting of the blades which occurs due to centrifugal force and to avoid harmful vibration. Accordingly, there is a large variation in blade performance from root to tip.

The first stage rotor blades in a compressor, in accordance with the invention, have a transonic airfoil section in order to obtain both high pressure ratio and high efficiency within a broad operating range. The relative entering velocity of the gas is preferably slightly supersonic, that is, at a Mach number of approximately 1.2, which is considered to be the upper end of the transonic range. Pressure r-a-tios can be obtained on the order of 1.6/1, or more, at efficiencies as high as 90%. The radial inlet to the transonic stage occurs at a very low velocity so that the usual complications with inlet ducts, strut fairin-gs, and the like in gas turbine engines are avoided. There is greater freedom also in the placement of accessory drives and power take-offs, making possible smaller and more compact gas turbines.

The centrifugal impeller in the compressor of the invention is shrouded, thereby eliminating tip leakages and improving flow conditions. Since the gas is turned radially outward again in the centrifugal stage, the combined transonic and centrifugal compressor rotor and easing can be made very short, thus contributing further to small size and compactness of the gas turbine. In view of the high efiiciency and pressure ratio of the transonic stage and of the improved efliciency of the shrouded centrifugal impeller, the overall pressure ratio and eficiency of the combined transonic and centrifugal stages are higher than can be obtained with a centrifugal stage alone.

It would sometimes be preferable, if it were possible, to go to higher relative entering blade Mach numbers in order to produce higher pressure ratios than can be obtained with transo-nic blades. Much work has been done in the past on the development of supersonic compressors, but no practical machines have ever been evolved. A primary difiiculty has resided in the diffusing process because of the threedimensional flow characteristics of the air leaving the rotating blade row. These characteristics have been aggravated by the twisting of the rotor blades due to the combined aerodynamic and centrifugal loads.

In a compressor, in accordance with the invention, however, the two-dimensional flow characteristics make it possible to avoid the last-mentioned difficulties previously encountered in providing diffusion from a supersonic rotor stage, so that, for the first time, it becomes feasible to make a practical application of supersonic compression. Further, inasmuch as the rotor blades are supported at each end, blades can be used which are thinner and therefore more efficient than those installed in axial flow machines without creating a danger of the blades twisting under load or vibrating.

Additionally, two compressor units, each comprising a transonic centripetal stage, a stator stage, and a centrifugal stage, can be connected together in series in order to obtain the high pressure ratios that are necessary in some gas turbine engines to achieve the low specific fuel consumptions required. Pressure ratios of 10-12/1 can be obtained in this manner with flow rates as low as approximately 1 lb./sec. and overall efficiencies of close to The two compressor units can be mounted on the same shaft and the tip diameters of the various stages selected to give the desired distribution of work between the individual stages. Air leaving the first centrifugal stage flows through a diffuser into a second :plenum chamber surrounding the second centripetal stage. The overall length of a two unit compressor is short, thereby making it possible to build small gas turbine engines that are relatively small in size, compact, and light weight, with low specific fuel consumptions.

Other objects and advantages of the invention will be apparent to those skilled in the art from the following detailed description of an exemplary embodiment of a compressor, taken in conjunction with the figures of the accompanying drawing, in which:

FIGURE 1 is an elevational view in section or" the compressor utilized in a gas turbine engine;

PEG. 2 is a partial view in section through the transonic rotor and stator blades taken generally along a plane through line 2-2 of FIG. 1;

FIG. 3 is a velocity vector diagram for the transonic rotor of FIG. 2;

FIG. 4- is an elevational view in section of a two unit compressor, in accordance with the invention.

FIG. 5 is a partial view in section through an inlet valve to the second plenum chamber of the two unit compressor; and

FIGS. 6A and 6B are, respectively, a partial view in section taken on a transverse plane through the two unit compressor of an inlet valve to the first plenum chamber and a partial elevational View in section through an inlet port to the first plenum chamber of the two unit compressor.

Referring to FIG. 1, a compressor casing 1 is provided with openings 2 through which air enters the compressors. The air flows first to an annular plenum chamber 3 and thence radially inward into a rotor 4 journaled in casing 1. Transonic rotor blades 5 on the rotor compress the air drawn from plenum chamber 3 and cause it to flow radially inward through stator blades 6, then axially through an annular passage 7 and into a centrifugal wheel 8, where it is further compressed. The air then flows radially outward through a difiuser 9 to the annular discharge passage 10. It will be noted that the compressor rotor 4 and casing 1 are very short.

The rotor blades 5, which have a transonic airfoil cross section, may best be made by cutting them from an extruded strip. The rotor blades 5 extend between two end rings 11 which are punched with slots matching the blade profile so that the blades can be inserted into them and brazed to the rings. The rings 11 fit into grooves in rings 12 to which they are also brazed or bonded. The (rings 12 fit into grooves in the opposed faces of discs 13 and 14. Studs 15 are butt welded or otherwise secured to rings 12 to permit bolting discs 13 and 14 to the rotor blade and ring assembly.

The stator blades 6 may likewise best be made by cutting them from an extruded strip and brazing them to the two side walls 16 and 17 which have previously been punched with slots matching the blade profile so that the blades can be inserted into them and brazed in a manner similar to the rotor blades. The stator blades 6 are designed to remove the whirl from the air so that the air flows radially inward with no tangential component. The stator blade assembly is brazed or welded to the flange member 18 which is bolted to casing 1 by means of bolts 19. Bearing cage 20 supporting shaft bearing 21 is clamped between flange 1 8 and easing 1 by bolts 19.

A labyrinth type seal 22 between disc 14 and stator passage wall 16 prevents the leakage of air from between the rotor and stator stages back to plenum chamber 3.

'Seal 22 consists of a series of concentric rings spaced a small distance apart from each other radially. Alternate rings are attached by brazing or other suitable means to the stator wall 16 or the disc 14, so that any leakage air would have to take a zig zag path through the seal. A similar seal 23 is used between stator wall 17 and disc 13 to prevent leakage of air from the annular passage 7 back to the entrance to the stator blades 6.

A portion of the impeller shroud 24 is cast integrally with impeller 8', or it may be brazed or otherwise integrally attached to the impeller. Spacer 25 is bolted to impeller shroud 24 and to disc 13 and thus provides the support for the t ransonic rotor stage comprising blades 5, rings 11 and 12, and discs 13 and 14. The inner portion 26 of spacer 25 serves both as a shroud for the inner portion of impeller 8 and as a pilot for disc 13. A labyrinth seal 27 is located on the outer diameter of spacer 25. Thin rings 28, formed on the spacer, engage the inner diameter of casing 1 to prevent leakage between the impeller discharge 29 and plenum chamber 3.

Referring to FIG. 3, air which has previously been drawn into the plenum chamber 3 through the slots 2 approaches the tips of the transonic blades 5 at a velocity nd in a radially inward direction as represented by the vector V inasmuch as the tips of the blades are travelling at a tangential velocity represented by the vector U the relative velocity of the air with respect to the blade is given by the vector V At the trailing edges of the blades 5 the absolute velocity of the air and the direction of flow are changed, as represented by the vector V in FIG. 3. The velocity of the blades is U at their trailing edges, and the relative velocity of the air with respect to the blades is V Referring now to FIG. 4, a two-unit compressor comprises two centripetakcentritugal compressors 50 and 50 of the type described above connected in series; that is, each consists of a centripetal stage of blades 54, 54', a stator stage of blades 56, 56 for removing the whirl in the air, a centrifugal stage 58, 58 and its diffuse-r 59, 59. Each centripetal stage has a plenum chamber 60, 60 from which it receives air. The air enters plenum chamber 60 through ports 61, while plenum chamber 60' receives its air through the annular discharge passage 63 from diffuser 59.

In some applications it may be desirable to have inlet ports to both plenum chambers, the port openings controlled by throttle valves. With this arrangement the second plenum chamber 60' could receive air either from the annular discharge passage 63 from the first unit, or directly from the atmosphere, depending on the throttle valve positions. FIG. 5 shows the port 62 and throttle valve 64' for plenum chamber 60', while FIG. 6 shows a port 62 and throttle valve 64 for plenum chamber 60. Normally, only one of the throttle valves would be open at a time while the other remained sealed tightly against its seat. Because of differences in air fiow and pressure ratio with the alternate throttle valve positions it is possible for a gas turbine engine with this arrangement to have a double rating, a normal or high power rating and a low power rating.

More particularly, maximum air flow and pressure ratio are obtained with the inlet valve 64 open to the plenum chamber 60 of the first compressor unit 50, while the inlet valve 64. to the second plenum chamber 60' remains closed. In this case, partially compressed air fiows from the first compressor unit 50 through difiusor 59 into the plenum chamber 60 of the second unit 5%) thence through the second unit which compresses it further, and discharges the air through the diffusor 59 to the gas turbine combustion chamber. On the other hand, when the first stage, or high power, valve 64 is closed and inlet air is admitted through the low power valve 64 directly to the second compressor unit 50', the air flow and pressure ratio are greatly reduced.

The throttling valves 64 and 64' make it possible for a single gas turbine engine to operate over a wide range of power outputs from full power down to very small percentages of full power. The high pressure ratio and compressor efiiciency enable the engine to have good fuel economy at high powers. At the same time, with the engine in its low power configuration, that is with valve 64 closed and 64 open, fuel consumption at. idle is also low because the power at idle is such a low percentage of the full power output. Windage losses in the first compressor unit 50 are held to a negligible amount when valve 64 is closed by sealing the valve tightly against its seat so that even a fast idle, which is required for automotive vehicles to obtain rapid acceleration, would be at a relatively low percentage of maximum power output, and the advantage of relatively low fuel consumption at idle could be combined with good acceleration ca ability, p

The valving' means for controlling the air flow and power output are not large or complicated, and, accordingly, they add little to engine cost or to the overall size and weight. The invention therefore makes possible small, relatively low cost gas turbine engines that would be suitable for use in compressed air units and many types of vehicular applications, especially those in which large amounts of time are spent at high power and at idle and very little time in between.

It will be understood by those skilled in the art that the above described embodiment of the compressor is merely exemplary and that it is susceptible of considerable variation and modification without departing from the spirit and scope of the invention. Such modifications and variations are intended to be within the scope of the appended claims.

I claim:

1. A compressor for compressing air and other elastic fluids, comprising an annular plenum chamber for receiving inlet air to the compressor, an annular row of transonic rotor blades communicating with said plenum chamber and arranged to receive and discharge air in a radially inward direction, an annular row of stator blades inwardly of said rotor blades, said stator blades having inner ends which extend radially inward so that the air leaves the stator blades with a velocity having substantially no tangential component, wall means defining an annular passage downstream of said stator blades and having an inner portion which is curved so that its downstream end lies generally axially and the air is directed axially as it leaves said passage, and centrifugal impeller means arranged to receive the air in an axial direction and discharge and compress the air in a radially outward direction.

2. A compressor as claimed in claim 1, wherein said stator blades are mounted between the outer ends of the walls defining said annular passage, said rotor blades are mounted between spaced-apart annular members secured to said centrifugal impeller means, and sealing means are provided between said annular members and the walls of said channel to limit the leakage of air discharged from said impeller means into the spaces between said channel walls and said rotor support members.

3. A compressor as claimed in claim 2, wherein said sealing means comprises a plurality of closely spaced concentric annular ribs on said annular members and a plurality of annular ribs on said walls and extending into the spaces between said ribs on said members to form a labyrinthic restrictor.

4. A compressor as claimed in claim 1, further comprising an annular row of spaced-apart radially arranged diffuser vanes downstream of said centrifugal impeller means.

5. A compressor for compressing air and other elastic fluids, comprising a centripetal first stage which includes a plurality of annularly arranged spaced-apart transonic blades, said blades mounted between spaced-apart annular members and arranged to direct and compress air inwardly, a stator stage which includes an annular row of blades having inner ends which curve radially inward so the air leaves the stator blades with a velocity having substantially no tangential component, said blades mounted between spaced-apart annular wall members having curved inner ends defining an annular passageway for turning the air discharged from the stator blades in an axial direction, labyrinth sealing means between the wall 6 members and the rotor blade mounting members, a cen trifugal stage which includes impeller means for compressing and directing the air outwardly and an external shroud forming at least a portion of an outer wall of the centrifugal stage, said centripetal stage secured to said centrifugal stage for rotation therewith.

6. A compressor for compressing air and other elastic fluids, comprising an annular plenum chamber for supplying inlet air to the compressor, an annularly disposed row of spaced-apart rotor blades communicating with said plenum chamber and arranged to receive the air and compress and discharge the air in a radially inward direction, an annularly disposed row of spaced-apart stator blades downstream of said rotor blades for receiving and directing the air in a radially inward direction with no tangential component, and impeller means downstream of said stator blades for receiving the air in an axial direction and discharging and compressing the air in an outward direction, means defining an annular discharge chamber downstream of said impeller means, and sealing means for preventing the communication of air between said plenum chamber and said outlet chamber, said sealing means including a plurality of radially arranged annular rings and a cylindrical member mounted outwardly of the rings and closely spaced from the rings which are expandable outwardly by thermal and centrifugal forces to eliminate any clearance gap between said discs and said cylindrical member.

7. A compressor for compressing air and other elastic fluids, comprising a first compressor unit and a second compressor unit, each of said compressor units having an annular plenum chamber for receiving inlet air, at least one row of transonic rotor blades communicating with said plenum chamber and arranged to receive the air and discharge and compress it in a radially inward direction, at least one row of stator blades disposed inwardly of said rotor blades and having inner ends which extend radially inward so that the air leaves the stator blades with a velocity having substantially no tangential component, means defining a passage inward of said stator blades for turning air leaving said stator blades to a substantially axial direction, centrifugal impeller means arranged to receive the air from the passage and discharge and compress the air in a radially outward direction, and a discharge passage downstream of said centrifugal impeller means, the discharge passage of the first compressor unit communicating with the plenum chamber of the second compressor unit.

References Cited by the Examiner UNITED STATES PATENTS 765,935 7/1904 Ray 103-97 1,213,889 1/1917 Lawaczeck 103-95 1,988,163 1/1935 Church 230-124 2,083,447 6/1937 Hoifmann 230-114 2,142,596 3/1939 Algarsson 230-114 2,265,806 12/1941 Goldschmied 103-95 2,429,978 11/1947 Blanchard 103-95 2,881,972 4/1959 Feilden 230-127 2,897,917 8/1959 Hunter 230-127 2,923,246 2/ 1960 Wright 103-95 2,925,952 2/ 1960 Garve 230- 2,985,108 5/1961 Stoner 103-95 3,156,407 11/ 1964 Bourguard 230-120 DONLEY I. STOCKING, Primary Examiner.

LAURENCE V. EFNER, HENRY F. RADUAZO,

MARTIN P. SCHWADRON, Examiners, 

1. A COMPRESSOR FOR COMPRESSING AIR AND OTHER ELASTIC FLUIDS, COMPRISING AN ANNULAR PLENUM CHAMBER FOR RECEIVING INLET AIR TO THE COMPRESSOR, AN ANNULAR ROW OF TRANSONIC ROTOR BLADES COMMUNICATING WITH SAID PLENUM CHAMBER AND ARRANGED TO RECEIVE AND DISCHARGE AIR IN A RADIALLY INWARD DIRECTION, AN ANNULAR ROW OF STATOR BLADES INWARDLY OF SAID ROTOR BLADES, SAID STATOR BLADES HAVING INNER ENDS WHICH EXTEND RADIALLY INWARD SO THAT THE AIR LEAVES THE STATOR BLADES WITH A VELOCITY HAVING SUBSTANTIALLY NO TANGENTIAL COMPONENT, WALL MEANS DEFINING AN ANNULAR PASSAGE DOWNSTREAM OF SAID STATOR BLADES AND HAVING AN INNER PORTION WHICH IS CURVED SO THAT ITS DOWNSTREAM END LIES GENERALLY AXIALLY AND THE AIR IS DIRECTED AXIALLY AS IT LEAVES SAID PASSAGE, AND CENTRIFUGAL IMPELLER MEANS ARRANGED TO RECEIVE THE AIR IN AN AXIAL DIRECTION AND DISCHARGE AND COMPRESS THE AIR IN A RADIALLY OUTWARD DIRECTION. 